A gas-blocking contactless seal for a shaft can comprise a sealing housing surrounding the shaft, a stator seal ring disposed in the housing and a rotor sealing ring mounted on the shaft.
The two sealing rings have juxtaposed end faces separated by the sealing gap in which the relative rotation of the rings generates a gas pressure sufficient to seal between the housing and the shaft.
The rotor sealing ring can be composed of a material of a high thermal conductivity and high modulus of elasticity as well as of high hardness. The sealing gap can be defined between the stator sealing ring and a rotor sealing ring so that a predetermined operating differential pressure is generated across the gap in the operating state of the seal, i.e. upon rotation of the shaft at its operating speed, and the stator sealing ring can be biased by a spring against the force of the sealing pressure while a rubber or plastic O-ring can seal between the stator sealing ring and a cylindrical portion of the sealing housing which extends coaxial with and parallel to the shaft.
With such a sealing arrangement it is important to distinguish between the functional annual gap, i.e. the gap between the cylindrical portion and the stator ring, and the mounting ring gap or tolerance, i.e. the gap or clearance which is provided to enable, for example, the stator ring to be mounted.
The functional gap is developed out of the mounting ring gap by a kind of displacement in the mounted condition by the effect of the sealing operating pressure differential.
It will be understood that there are various constructions of prior art systems which utilize this approach to seal around a shaft. Some of these systems date back to 1925 and a variety of materials having the requisite hardness for such sealing systems is described, for example, in VDI Zeitschrift 102 (1960) No. 18 pages 728 to 732.
In one sealing arrangement of the aforedescribed type (see European Patent Publication EP 0013678) the recesses of the sealing rings which generate the pressure gradient are spiral grooves which extend from a peripheral edge of the respective sealing ring. In this case, only the rotor sealing ring is composed of a material of high thermal conductivity, modulus of elasticity and hardness.
The stator sealing ring is composed of a material of relatively low modulus of elasticity and minimal hardness, for example, of carbon whose thermal conductivity is not significant.
Because of the relatively low modulus of elasticity and thermal conductivity of the thermal stator ring, this prior art sealing arrangement can give rise to a torsional deformation (twist) of the sealing assembly at least partially as a consequence of the operating temperature and this, in turn, can result in a cocking of one of the rings.
The temperature gradient in the axial direction can in such cases be 25.degree. C. or more. The torsional deformation of the stator sealing ring can detrimentally effect the seal formed during operation, can result in contact at inappropriate times between the sealing rings and can be disadvantageous for a variety of other reasons.
It has been proposed to overcome the problem by counteracting the torsional deformation moment, twist or torque by a moment or torque generated by an appropriate pressure distribution in the sealing gap so that the two torques are effective in opposite senses. For this purpose, the recesses are formed as spiral grooves which are designed to have a pumping effect at least for the spiral grooves formed in the rotor sealing ring and these grooves can extend toward the opposite periphery from one of the peripheries of the annular sealing surface. These grooves can end in a dam or rib so that the groove is not throughgoing from one peripheral edge to the other. As a function of the spiral groove depth and the inter groove rib width as well as a function of the equilibrium between the various parameter, it is possible to create the torque balance desired wherein the torsion effect is counterbalanced by the torque or moment generated by the gas pressure.
Even when these conditions are scrupulously observed, the desired effect cannot always be achieved or the effect which is obtained can be unsatisfactory. For example, the described equilibrium cannot be obtained under all operating conditions. In practice, the plane parallelity of the confronting sealing surface can be restored only to a maximum of about 70%.
The problem appears to be that conventional systems are not adequately able to take into consideration significant tribological characteristics of the seal, like inertia, stiffness, frictional moment etc. As a consequence, upon resetting of the torsional deformation or twist, the deformation may not be sufficiently compensated.
The most readily observable result of these problems with conventional systems is a detrimental leakage rate which tends to increase with increasing speed of the shaft and thus with increasing speeds of the rotor sealing ring because of the pumping effect of the spiral grooves. This leakage phenomenon tends to increase even more as a result of the incomplete restoration of the shape if a sealing ring following a torsional deformation as described.
The fact that, in the prior art system, a temperature-dependent and differential-pressure-dependent torsional deformation of the stator sealing ring is permissible and is in part caused by the arrangement of the polar moment of inertia of the stator sealing ring, so that the described pressure distribution in the gap is necessary to reset it, has a further drawback which is detrimental to the effective operation of the seal. The described twist requires, in accordance with the laws of mechanics, that the mounting ring gap be significantly greater than the functional annular gap formed when the stator sealing ring has undergone twisting and without consideration of the resetting of the twist resulting from the pressure distribution in the pressurized gas gap. The dimensions of the mounting ring gap (mounting tolerance) and the ratio of the mounting ring gap to the functional gap under the various operating conditions cannot be related to the resetting in most instances because the temperature dependent deformation and the resetting which is dependent upon the operating state of the sealing arrangement do not occur simultaneously and both must occur under conditions that a contact of the sealing surface of the rotor and stator sealing rings does not occur.
In practice, under all conventional diameter relationships of the sealing arrangement in the conventional system, the width of the mounting ring gap is in the range of 0.4 to 0.5 mm and usually, therefore, is in excess of 0.4 mm.
That has meant that O-ring of the conventional device should have a corresponding hardness, for example, a Shore A hardness of 90 or more in accordance with German Industrial Standards DIN 53505.
In spite of this adaptation of the O-ring to the conditions described, it has been found that operating pressure differentials above 80 bar or reaching a maximum of 100 bar cannot be provided in a practical sense if the sealing assembly is to be used for the conventional operating life of several thousand hours.
On the other hand, the industrial need is increasing for sealing arrangements which require much higher operating pressure differentials, i.e. pressure differentials which can be well above 100 bar. There is another aspect of the operation of a conventional sealing of the type described which should be mentioned. If one applies an O-ring of high hardness ahead of the annular gap between the stator sealing ring and the cylindrical portion of the sealing housing, the latter is subjected to nonuniform stresses resulting from the fact that the relative rotation of the shaft with the rotor sealing ring and the surrounding housing is neither perfectly round nor perfectly coaxial with the stator sealing ring and the associated housing parts. Accordingly, vibrations can result as induced by the stator sealing ring. These vibrations apply to the O-ring.
The functional annular gap must be defined with a certain amount of play to accommodate these vibrations without detriment to the sealing effectiveness. When the O-ring is composed of a very hard material it is not generally able to follow the induced movements of the stator sealing ring and thus may be unduly stressed and can become damaged. It may, if hard enough, limit the play required as described above, resulting in a detriment to the sealing effectiveness. In particular, this can lead to detrimental contact between the sealing end faces of the rotor sealing ring and the stator sealing ring.